JP6319336B2 - Engine oil supply device - Google Patents

Description

  The present invention relates to an engine oil supply apparatus.

  2. Description of the Related Art Conventionally, an oil supply device that supplies oil to each part of an engine is known.

  In Patent Document 1, the amount of oil supplied to the crank bearing or the like by the flow rate adjustment valve is limited so that the oil pressure of the variable valve mechanism becomes equal to or higher than the target pressure when the engine is started. A technique for ensuring operability is disclosed.

  Patent Document 2 includes a plurality of hydraulic actuators (hydraulic VVT, valve stop device), sets a target hydraulic pressure to the maximum required hydraulic pressure among the required hydraulic pressures of the plurality of hydraulic actuators according to the operating state of the engine, Disclosed is a technique for controlling the discharge amount of a variable displacement oil pump with an oil control valve so that the actual hydraulic pressure matches the set target hydraulic pressure.

Japanese Patent No. 5168372 JP 2014-199011 A

  When the engine is cold started, the oil pump operates with the oil viscosity high, so it is difficult to spread the oil throughout the oil supply path. For this reason, oil is not sufficiently supplied to the pressure chamber of the oil pump provided in the oil supply passage downstream of the main gallery, but oil is sufficiently supplied to the main gallery provided with the hydraulic sensor. Therefore, a situation occurs in which an excessively large actual oil pressure is detected by the oil pressure sensor.

  At this time, even if the oil control valve is controlled to reduce the actual oil pressure to the target oil pressure by feedback control, the oil pump cannot adjust the oil discharge amount because sufficient oil is not supplied to the pressure chamber. The actual hydraulic pressure cannot be lowered. Therefore, the control amount in the direction of decreasing the actual hydraulic pressure further increases.

  Eventually, sufficient oil is supplied to the pressure chamber, and the oil pump is in a state where the discharge amount can be adjusted, but at this time, the control amount in the lowering direction is excessively large, so the actual hydraulic pressure decreases at once, The target oil pressure is significantly below, and the oil pump discharge rate becomes excessively small.

  Since the amount of oil supplied to the pressure chamber depends on the discharge amount of the oil pump, if the discharge amount of the oil pump becomes excessively small, sufficient oil is not supplied to the pressure chamber, and the oil pump cannot be adjusted again. It becomes possible. Therefore, even if the oil control valve is controlled in the direction to raise the excessively low actual oil pressure to the target oil pressure, the actual oil pressure is not easily raised to the target oil pressure. Therefore, the control amount in the direction of increasing the actual hydraulic pressure increases. Eventually, the oil pump becomes controllable. At this time, however, the actual hydraulic pressure increases rapidly because the control amount in the direction of increasing the actual hydraulic pressure is excessive.

  Thus, when feedback control is applied at the time of cold start, there is a problem that the hunting of the actual hydraulic pressure increases and the time until the actual hydraulic pressure converges to the target hydraulic pressure is prolonged. As a result, there arises a problem that stable control of a hydraulic actuator such as the hydraulic VVT cannot be performed at an early stage.

  In addition, since the above Patent Documents 1 and 2 are all premised on feedback control, such a problem at the time of cold start cannot be solved.

  An object of the present invention is to provide an engine oil supply that can shorten the period until the actual hydraulic pressure converges to the target hydraulic pressure even during a cold start, and can perform stable control of the hydraulic actuator early. It is to provide a control device.

One embodiment of the present invention includes a variable displacement oil pump,
A hydraulic actuator connected to the oil pump via an oil supply path;
A hydraulic pressure detection unit for detecting the hydraulic pressure of the oil supply path;
An oil control valve that changes a discharge amount of oil discharged from the oil pump by controlling a flow rate of oil supplied to the pressure chamber of the oil pump;
A target hydraulic pressure is set from the required hydraulic pressure of the hydraulic actuator according to the operating state of the engine, and the oil control valve is controlled so that the actual hydraulic pressure detected by the hydraulic pressure detection unit matches the target hydraulic pressure, A control unit that feedback controls the discharge amount of the oil pump,
The control unit performs the feedback control after executing a fixed value control for setting the control value of the oil control valve to a fixed control value for a certain period from the start of the engine ,
The fixed control value indicates that the actual hydraulic pressure does not exceed a predetermined upper limit hydraulic pressure when the oil control valve is configured with an element having the largest variation among the variations in characteristics of the elements constituting the oil control valve. Under the condition, a value that matches or is close to the control value corresponding to the required hydraulic pressure of the hydraulic actuator at the start of the feedback control is set.

According to this aspect, the fixed value control is performed in which the control value of the oil control valve is set to the fixed control value for a certain period after the engine is started. For this reason, even during cold start when the oil viscosity is high, accumulation of the control value by the I term is eliminated as in the case of performing PID feedback control, and the actual hydraulic pressure can be quickly converged to the target hydraulic pressure. As a result, early stable control of the hydraulic actuator can be achieved.
Further, according to this aspect, the fixed control value is set to a value that matches or is close to the control value corresponding to the required hydraulic pressure of the hydraulic actuator at the start of feedback control, so control from fixed value control to feedback control is performed. Can be made early. Furthermore, the fixed control value is a condition that the actual hydraulic pressure does not exceed a predetermined upper limit hydraulic pressure when the oil control valve is configured with an element having the largest variation among the variations in characteristics of the elements constituting the oil control valve. Meet. Therefore, in the fixed value control, the actual hydraulic pressure does not exceed the upper limit hydraulic pressure, the reliability of the oil supply device is satisfied, and noise caused by the oil pump can be suppressed. Furthermore, since the actual hydraulic pressure does not exceed the upper limit hydraulic pressure, the engine is cooled by injecting oil from a hydraulic actuator such as an oil jet, and a decrease in combustion stability of the engine can be suppressed.

In the above aspect, the apparatus further comprises a viscosity characteristic detection unit that detects the viscosity characteristic of the oil,
The controller may set the fixed control value larger as the viscosity indicated by the viscosity characteristic detected by the viscosity characteristic detector decreases.

  According to this aspect, since the fixed control value is set to be larger as the viscosity indicated by the viscosity characteristic of the oil becomes smaller, the viscosity characteristic of the oil after the engine is started is taken into consideration and the actual hydraulic pressure is changed to the target hydraulic pressure. Convergence can be accelerated.

In the above aspect, the apparatus further comprises a viscosity characteristic detection unit that detects the viscosity characteristic of the oil,
The controller may set the fixed period shorter as the viscosity indicated by the viscosity characteristic detected by the viscosity characteristic detector decreases.

  According to this aspect, since the execution time of the fixed value control is set shorter as the viscosity indicated by the viscosity characteristic of the oil becomes smaller, the target of the actual hydraulic pressure is set in consideration of the viscosity characteristic of the oil after the engine is started. Convergence to hydraulic pressure can be accelerated.

In the above aspect, the apparatus further comprises a viscosity characteristic detection unit that detects the viscosity characteristic of the oil,
The control unit may execute the fixed value control when the viscosity detected by the viscosity characteristic detection unit is higher than a predetermined viscosity.

  According to this aspect, the fixed value control is executed only when the viscosity detected by the viscosity characteristic detection unit is higher than the predetermined viscosity. Therefore, when the viscosity is lower than the predetermined viscosity, the fixed value control is executed, and on the contrary, it can be prevented that the time until the actual hydraulic pressure converges to the target hydraulic pressure is increased.

  According to the present invention, at the time of cold start where the oil viscosity is high, the actual hydraulic pressure can be quickly converged to the target hydraulic pressure, and early stable control of the hydraulic actuator can be achieved.

1 is a schematic cross-sectional view of an engine cut along a plane including an axis of a cylinder. It is a longitudinal cross-sectional view of a crankshaft. It is sectional drawing which shows the structure and operation | movement of a hydraulically operated valve stop apparatus, (A) shows a locked state, (B) shows a lock release state, (C) shows the state which the action | operation of the valve has stopped. It is sectional drawing which shows schematic structure of a variable valve timing mechanism. It is a hydraulic circuit diagram of an oil supply apparatus. It is a graph which shows the time transition of duty ratio at the time of performing PID feedback control at the time of cold start, and real oil pressure. It is the graph which showed the characteristic of the oil control valve. It is a graph which shows the time transition of a duty ratio at the time of performing cold duty start and fixed duty control, and real oil pressure. It is a flowchart which shows an example of operation | movement of the oil supply apparatus in embodiment of this invention. The upper diagram shows an example of a fixed duty ratio table, and the lower diagram shows an example of a duration table. It is a figure which shows an example of the request | requirement hydraulic pressure table which shows the request | requirement hydraulic pressure of the exhaust side VVT.

  Hereinafter, exemplary embodiments will be described in detail with reference to the drawings.

  FIG. 1 is a schematic cross-sectional view of engine 100 cut along a plane including the axis of the cylinder. In the present specification, for convenience of explanation, the axial direction of the cylinder is referred to as the vertical direction, and the cylinder row direction is referred to as the front-rear direction. Further, in the cylinder row direction, the non-transmission side of engine 100 is referred to as a front side, and the transmission side is referred to as a rear side.

  The engine 100 is an in-line four-cylinder engine in which four cylinders are arranged in a predetermined cylinder row direction. The engine 100 includes a cylinder head 1, a cylinder block 2 attached to the cylinder head 1, and an oil pan 3 attached to the cylinder block 2.

  The cylinder block 2 has an upper block 21 and a lower block 22. The lower block 22 is attached to the lower surface of the upper block 21. An oil pan 3 is attached to the lower surface of the lower block 22.

  In the upper block 21, four cylinder bores 23 corresponding to the four cylinders are formed side by side in the cylinder row direction (only one cylinder bore 23 is shown in FIG. 1). The cylinder bore 23 is formed in the upper part of the upper block 21, and the lower part of the upper block 21 defines a part of the crank chamber. A piston 24 is inserted into the cylinder bore 23. The piston 24 is connected to the crankshaft 26 via a connecting rod 25. A combustion chamber 27 is defined by the cylinder bore 23, the piston 24, and the cylinder head 1. The four cylinder bores 23 correspond to a first cylinder, a second cylinder, a third cylinder, and a fourth cylinder in order from the front side.

  The cylinder head 1 is provided with an intake port 11 and an exhaust port 12 that open to the combustion chamber 27, and the intake port 11 is provided with an intake valve 13 that opens and closes the intake port 11. The exhaust port 12 is provided with an exhaust valve 14 that opens and closes the exhaust port 12. The intake valve 13 and the exhaust valve 14 are driven by cam portions 41a and 42a provided on the cam shafts 41 and 42, respectively.

  Specifically, the intake valve 13 and the exhaust valve 14 are urged in the closing direction (upward in FIG. 1) by the valve springs 15 and 16. Swing arms 43 and 44 are interposed between the intake valve 13 and the exhaust valve 14 and the cam portions 41a and 42a, respectively. One end portions of the swing arms 43 and 44 are respectively supported by hydraulic lash adjusters (hereinafter referred to as “HLA”) 45 and 46. The swing arms 43 and 44 swing around the one end portions supported by the HLA 45 and 46 as the cam followers 43a and 44a provided at substantially central portions thereof are pushed by the cam portions 41a and 42a, respectively. By swinging in this way, the swing arms 43 and 44 move the intake valve 13 and the exhaust valve 14 at the other end in the opening direction (downward in FIG. 1) against the urging force of the valve springs 15 and 16, respectively. Let The HLA 45 and 46 automatically adjust the valve clearance to zero by hydraulic pressure.

  The HLA 45 and 46 provided in the first cylinder and the fourth cylinder are provided with valve stop mechanisms that stop the operations of the intake valve 13 and the exhaust valve 14, respectively. Hereinafter, when HLA is distinguished by the presence or absence of a valve stop mechanism, the HLA 45 and 46 having the valve stop mechanism are referred to as HLA 45a and 46a, and the HLA 45 and 46 having no valve stop mechanism are referred to as HLA 45b and 46b. Called. The engine 100 operates all the intake valves 13 and the exhaust valves 14 of the first to fourth cylinders during all cylinder operation, while the intake valves 13 and the exhaust valves 14 of the first and fourth cylinders operate during the reduced cylinder operation. The operation is stopped, and the intake valve 13 and the exhaust valve 14 of the second and third cylinders are operated.

  In portions corresponding to the first and fourth cylinders of the cylinder head 1, mounting holes for mounting the HLA 45a and 46a are formed. The HLA 45a and 46a are mounted in the mounting hole. The cylinder head 1 is formed with an oil supply passage communicating with the mounting hole. Oil is supplied to the HLA 45a and 46a through this oil supply passage.

  A cam cap 47 is attached to the upper part of the cylinder head 1. The cam shafts 41 and 42 are rotatably supported by the cylinder head 1 and the cam cap 47.

  An intake side oil shower 48 is provided above the intake side camshaft 41, and an exhaust side oil shower 49 is provided above the exhaust side camshaft 42. The oil shower 48 on the intake side and the oil shower 49 on the exhaust side are configured to drop oil on contact portions between the cam portions 41 a and 42 a and the cam followers 43 a and 44 a of the swing arms 43 and 44.

  The engine 100 is also provided with a variable valve timing mechanism (hereinafter referred to as “VVT”) that changes the valve characteristics of the intake valve 13 and the exhaust valve 14. The intake side VVT is an electric type, and the exhaust side VVT 18 is a hydraulic type.

  The upper block 21 has a first side wall 21 a located on the intake side with respect to the four cylinder bores 23, a second side wall 21 b located on the exhaust side with respect to the four cylinder bores 23, and a front side relative to the frontmost cylinder bore 23. A plurality of vertical walls extending in the vertical direction at a portion between a front wall (not shown) positioned, a rear wall (not shown) positioned rearward of the rearmost cylinder bore 23, and two adjacent cylinder bores 23 Wall 21c.

  The lower block 22 corresponds to the first side wall 21a of the upper block 21, and corresponds to the first side wall 22a located on the intake side, and the second side wall 22b located on the exhaust side corresponding to the second side wall 21b of the upper block 21. The front wall corresponding to the front wall of the upper block 21 (not shown), the rear wall corresponding to the rear wall of the upper block 21 and not shown, and the upper block 21 It has the some vertical wall 22c corresponding to the vertical wall 21c. The upper block 21 and the lower block 22 are bolted.

  There is a crankshaft 26 between the front wall of the upper block 21 and the front wall of the lower block 22, between the rear wall of the upper block 21 and the rear wall of the lower block 22, and between the vertical wall 21c and the vertical wall 22c. The bearing part 28 which supports is provided.

  Hereinafter, the bearing portion 28 between the vertical wall 21c and the vertical wall 22c will be described with reference to FIG. FIG. 2 is a cross-sectional view of the vertical wall 21c of the upper block 21 and the vertical wall 22c of the lower block 22 located at the center in the cylinder row direction. Similar bearings 28 are provided between the front wall of the upper block 21 and the front wall of the lower block 22 and between the rear wall of the upper block 21 and the rear wall of the lower block 22. When distinguishing each bearing part 28, it calls the 1st bearing part 28A, the 2nd bearing part 28B, the 3rd bearing part 28C, the 4th bearing part 28D, and the 5th bearing part 28E sequentially from the front side.

The bearing portion 28 is provided between two bolt fastening locations. Specifically, the bearing portion 28 is disposed between the pair of screw holes 21f and the bolt insertion holes 22f. The bearing portion 28 has a cylindrical bearing metal 29. A semicircular cutout is formed at each joint between the vertical wall 21c and the vertical wall 22c. The bearing metal 29 has a divided structure composed of a first semicircular portion 29a and a second semicircular portion 29b, and the first semicircular portion 29a is attached to a notch portion of the vertical wall 21c, and the second semicircular portion is provided. The portion 29b is attached to the cutout portion of the vertical wall 22c. By combining the vertical wall 21c and the vertical wall 22c, the first semicircular portion 29a and the second semicircular portion 29b are combined to form a cylindrical shape. An oil groove 29c extending in the circumferential direction is formed on the inner peripheral surface of the first semicircular portion 29a. In addition, the first semicircular portion 29a is formed with a connecting passage 29d having one end opened to the outer peripheral surface of the first semicircular portion 29a and the other end opened to the oil groove 29c. An oil supply passage is formed in the upper block 21, and oil is supplied to the outer peripheral surface of the first semicircular portion 29a through the oil supply passage. The communication path 29d is disposed at a position communicating with the oil supply path. As a result, the oil supplied from the oil supply passage flows into the oil groove 29c through the communication passage 29d.

  Although not shown, a chain cover is attached to the front wall of the cylinder block 2. A drive sprocket provided on the crankshaft 26, a timing chain wound around the drive sprocket, a chain tensioner for applying tension to the timing chain, and the like are disposed inside the chain cover.

  The HLA 45a and 46a provided with a valve stop mechanism will be described in detail with reference to FIG. Since the configurations of the HLA 45a and 46a are substantially the same, only the HLA 45a will be described below.

  The HLA 45a has a pivot mechanism 45c and a valve stop mechanism 45d.

  The pivot mechanism 45c is a known HLA pivot mechanism, and automatically adjusts the valve clearance to zero by hydraulic pressure. The HLA 45b and 46b do not have a valve stop mechanism, but have substantially the same pivot mechanism as the pivot mechanism 45c.

  The valve stop mechanism 45d is a mechanism that switches between operation and stop of the corresponding intake valve 13 or exhaust valve 14. The valve stop mechanism 45d has an opening at one end, a bottom at the other end, and an outer cylinder 45e that accommodates the pivot mechanism 45c so as to be slidable in the axial direction, and is opposed to the side peripheral surface of the outer cylinder 45e. A pair of lock pins 45g inserted in the two through holes 45f so as to be able to advance and retreat, a lock spring 45h for urging one lock pin 45g outward in the radial direction of the outer cylinder 45e, a bottom of the outer cylinder 45e and a pivot A lost motion spring 45i is provided between the mechanism 45c and urges the pivot mechanism 45c in the axial direction toward the opening of the outer cylinder 45e. The lock pin 45g is disposed at the lower end of the pivot mechanism 45c. The lock pin 45g is driven by hydraulic pressure, and is switched between a state in which the lock pin 45g is fitted in the through hole 45f and a state in which the engagement with the through hole 45f is released by moving inward in the radial direction of the outer cylinder 45e.

  As shown in FIG. 3A, when the lock pin 45g is fitted in the through hole 45f, the pivot mechanism 45c protrudes from the outer cylinder 45e with a relatively large protrusion amount, and the lock pin 45g causes the outer cylinder 45e to move. Movement in the axial direction is restricted. That is, the pivot mechanism 45c is in a locked state. In this state, the top of the pivot mechanism 45c contacts one end of the swing arm 43 or the swing arm 44, and functions as a fulcrum for swinging. Accordingly, the swing arms 43 and 44 move the intake valve 13 and the exhaust valve 14 in the opening direction against the urging force of the valve springs 15 and 16 at the other ends, respectively. That is, when the valve stop mechanism 45d is in the locked state, the corresponding intake valve 13 or exhaust valve 14 is operated.

  On the other hand, when hydraulic pressure acts on the lock pin 45g from the outside in the radial direction, the lock pin 45g moves inward in the radial direction of the outer cylinder 45e against the urging force of the lock spring 45h, as shown in FIG. Then, the fitting with the through hole 45f is released. Thereby, the lock of the pivot mechanism 45c is released.

  Even in this unlocked state, the pivot mechanism 45c protrudes from the outer cylinder 45e with a relatively large protrusion amount by the urging force of the lost motion spring 45i. However, the pivot mechanism 45c is not restricted from moving in the axial direction of the outer cylinder 45e, and can move. The urging force of the lost motion spring 45i is set to be smaller than the urging force of the valve springs 15 and 16 for urging the intake valve 13 and the exhaust valve 14 in the closing direction. Therefore, when the cam followers 43a and 44a are pushed by the cam portions 41a and 42a in the unlocked state, the top portions of the intake valve 13 and the exhaust valve 14 become fulcrums of the swing arms 43 and 44, respectively. 3C, the pivot mechanism 45c is moved toward the bottom of the outer cylinder 45e against the urging force of the lost motion spring 45i. That is, when the valve stop mechanism 45d is in the unlocked state, the operation of the corresponding intake valve 13 or exhaust valve 14 is stopped.

  Next, the exhaust side VVT 18 will be described in detail with reference to FIG.

  The exhaust side VVT 18 includes a substantially annular housing 18a and a rotor 18b accommodated in the housing 18a. The housing 18a is connected to a cam pulley 18c that rotates in synchronization with the crankshaft 26 so as to be integrally rotatable. The rotor 18b is connected to a camshaft 41 that opens and closes the intake valve 13 so as to be integrally rotatable. The rotor 18b is provided with a vane 18d that slides with the inner peripheral surface of the housing 18a. A plurality of retarded hydraulic chambers 18e and advanced hydraulic chambers 18f defined by the inner peripheral surface of the housing 18a, the vanes 18d and the main body of the rotor 18b are formed in the housing 18a. Oil is supplied to the retard hydraulic chamber 18e and the advance hydraulic chamber 18f. When the hydraulic pressure in the retarded hydraulic chamber 18e is high, the rotor 18b rotates in the opposite direction with respect to the rotation direction of the housing 18a. That is, the camshaft 41 rotates in the opposite direction with respect to the cam pulley 18c, and the valve opening timing of the intake valve 13 is delayed. On the other hand, when the hydraulic pressure in the advance hydraulic chamber 18f is high, the rotor 18b rotates in the same direction with respect to the rotation direction of the housing 18a. That is, the camshaft 41 rotates in the same direction with respect to the cam pulley 18c, and the valve opening timing of the intake valve 13 is advanced.

  Next, the oil supply apparatus 200 will be described with reference to FIG. FIG. 5 shows a hydraulic circuit diagram of the oil supply apparatus 200.

  The oil supply apparatus 200 includes a variable capacity oil pump 81 that is rotationally driven by the crankshaft 26, and an oil supply passage 5 that is connected to the oil pump 81 and through which oil flows. Oil pump 81 is an auxiliary machine driven by engine 100.

  The oil pump 81 is a known variable displacement oil pump and is driven by the crankshaft 26. The oil pump 81 is attached to the lower surface of the lower block 22 and is housed in the oil pan 3. Specifically, the oil pump 81 includes a drive shaft 81a that is rotationally driven by the crankshaft 26, a rotor 81b that is coupled to the drive shaft 81a, and a plurality of vanes 81c that are capable of moving forward and backward in the radial direction from the rotor 81b. The cam ring 81d configured to accommodate the rotor 81b and the vane 81c and adjust the amount of eccentricity with respect to the rotation center of the rotor 81b, and bias the cam ring 81d in a direction in which the amount of eccentricity with respect to the rotation center of the rotor 81b increases. It includes a spring 81e, a ring member 81f disposed inside the rotor 81b, and a housing 81g that houses the rotor 81b, vane 81c, cam ring 81d, spring 81e, and ring member 81f.

  Although not shown, one end of the drive shaft 81a protrudes outward from the housing 81g, and a driven sprocket is connected to the one end. A timing chain is wound around the driven sprocket. This timing chain is also wound around the drive sprocket of the crankshaft 26. Thus, the rotor 81b is rotationally driven by the crankshaft 26 via the timing chain.

  When the rotor 81b rotates, each vane 81c slides on the inner peripheral surface of the cam ring 81d. Accordingly, the pump chamber (hydraulic oil chamber) 81i is defined by the rotor 81b, the two adjacent vanes 81c, the cam ring 81d, and the housing 81g.

The housing 81g is formed with a suction port 81j through which oil is sucked into the pump chamber 81i and a discharge port 81k through which oil is discharged from the pump chamber 81i. An oil strainer 81l is connected to the suction port 81j. The oil strainer 81 l is immersed in the oil stored in the oil pan 3. That is, the oil stored in the oil pan 3 is sucked into the pump chamber 81i from the suction port 81j through the oil strainer 81l. On the other hand, the oil supply path 5 is connected to the discharge port 81k. That is, the oil boosted by the oil pump 81 is discharged from the discharge port 81k to the oil supply passage 5.

  The cam ring 81d is supported by the housing 81g so as to swing around a predetermined fulcrum. The spring 81e biases the cam ring 81d toward one side around the fulcrum. A pressure chamber 81m is defined between the cam ring 81d and the housing 81g. The pressure chamber 81m is configured to be supplied with oil from the outside. The oil pressure in the oil pressure chamber 81m acts on the cam ring 81d. Therefore, the cam ring 81d swings according to the balance between the biasing force of the spring 81e and the hydraulic pressure of the pressure chamber 81m, and the amount of eccentricity of the cam ring 81d with respect to the rotation center of the rotor 81b is determined. The capacity of the oil pump 81 changes according to the amount of eccentricity of the cam ring 81d, and the amount of oil discharged changes.

  The oil supply path 5 is formed by a pipe and a flow path formed in the cylinder head 1 and the cylinder block 2. The oil supply passage 5 includes a main gallery 50 extending in the cylinder row direction in the cylinder block 2, a first communication passage 51 connecting the oil pump 81 and the main gallery 50, and a second communication passage extending from the main gallery 50 to the cylinder head 1. 52, a third communication passage 53 extending in a substantially horizontal direction between the intake side and the exhaust side in the cylinder head 1, a control oil supply passage 54 branched from the first communication passage 51, and a branch from the third communication passage 53 And first to fifth oil supply passages 55 to 59.

  The first communication path 51 is connected to the discharge port 81 k of the oil pump 81. In the first communication passage 51, an oil filter 82 and an oil cooler 83 are provided in order from the oil pump 81 side. That is, the oil discharged from the oil pump 81 to the first communication passage 51 is filtered by the oil filter 82, the oil temperature is adjusted by the oil cooler 83, and then flows into the main gallery 50.

  In the main gallery 50, an oil jet 71 that injects oil to the back side of the four pistons 24, a bearing metal 29 of five bearing portions 28 that rotatably supports the crankshaft 26, and four connecting rods 25 rotate. A bearing metal 72 disposed on a freely connected crank pin, an oil supply part 73 that supplies oil to the hydraulic chain tensioner, an oil jet 74 that injects oil to the timing chain, and oil that circulates through the main gallery 50 A hydraulic pressure sensor 50a for detecting the hydraulic pressure is connected. The hydraulic pressure sensor 50a is an example of a hydraulic pressure detection unit. Oil is always supplied to the main gallery 50. The oil jet 71 has a check valve and a nozzle, and when a hydraulic pressure of a predetermined value or more acts, the check valve opens to inject oil from the nozzle.

  Further, from the main gallery 50, a control oil supply passage 54 connected to the pressure chamber 81 m of the oil pump 81 is branched via an oil control valve 84. An oil filter 54 a is provided in the control oil supply passage 54. The oil in the main gallery 50 passes through the control oil supply passage 54, the oil pressure is adjusted by the oil control valve 84, and then flows into the pressure chamber 81 m of the oil pump 81. That is, the oil control valve 84 changes the discharge amount of the oil pump 81 by controlling the flow rate of the oil supplied to the pressure chamber 81m.

  The oil control valve 84 is a linear solenoid valve. The oil control valve 84 adjusts the flow rate of the oil supplied to the pressure chamber 81m according to the input duty ratio.

  The second communication path 52 allows the main gallery 50 and the third communication path 53 to communicate with each other. Oil flowing through the main gallery 50 flows into the third communication path 53 through the second communication path 52. The oil flowing into the third communication path 53 is distributed to the intake side and the exhaust side of the cylinder head 1 through the third communication path 53.

  The first oil supply passage 55 includes an oil supply portion 91 for a bearing metal that supports the cam journal of the intake camshaft 41, an oil supply portion 92 for a thrust bearing of the intake camshaft 41, and an HLA 45a with a valve stop mechanism. The pivot mechanism 45c, the HLA 45b without a valve stop mechanism, the oil shower 48 on the intake side, and the oil supply portion 93 of the sliding portion of the intake side VVT are connected.

  The second oil supply passage 56 includes a bearing metal oil supply portion 94 that supports the cam journal of the exhaust camshaft 42, a thrust bearing oil supply portion 95 of the exhaust camshaft 42, and an HLA 46a with a valve stop mechanism. The pivot mechanism 46c, the HLA 46b without a valve stop mechanism, and the oil shower 49 on the exhaust side are connected.

  The third oil supply passage 57 is connected to the retard hydraulic chamber 18e and the advance hydraulic chamber 18f of the exhaust side VVT 18 via the first direction switching valve 96. The third oil supply passage 57 is connected to an oil supply portion 94 located at the forefront portion of the oil supply portion 94 of the bearing metal of the camshaft 42 on the exhaust side. An oil filter 57 a is connected to the upstream side of the first direction switching valve 96 in the third oil supply passage 57. The flow rate of oil supplied to the retard hydraulic chamber 18e and the advance hydraulic chamber 18f is adjusted by the first direction switching valve 96.

  The fourth oil supply path 58 is connected to the valve stop mechanism 45d of the HLA 45a with the valve stop mechanism of the first cylinder and the valve stop mechanism 46d of the HLA 46a with the valve stop mechanism via the second direction switching valve 97. An oil filter 58 a is connected to the upstream side of the second direction switching valve 97 in the fourth oil supply path 58. The oil supply to the valve stop mechanism 45d and the valve stop mechanism 46d of the first cylinder is controlled by the second direction switching valve 97.

  The fifth oil supply passage 59 is connected via a third direction switching valve 98 to the valve stop mechanism 45d of the HLA 45a with a valve stop mechanism of the fourth cylinder and the valve stop mechanism 46d of the HLA 46a with a valve stop mechanism. An oil filter 59 a is connected to the upstream side of the third direction switching valve 98 in the fifth oil supply passage 59. The third direction switching valve 98 controls oil supply to the valve stop mechanism 45d and the valve stop mechanism 46d of the fourth cylinder.

  The oil supplied to each part of the engine 100 is dropped into the oil pan 3 through a drain oil passage (not shown) and is recirculated by the oil pump 81.

  Engine 100 is controlled by controller 60. The controller 60 has a processor and a memory, and receives detection results from various sensors that detect the operating state of the engine 100. For example, the controller 60 includes a hydraulic pressure sensor 50a, a crank angle sensor 61 that detects the rotation angle of the crankshaft 26, an air flow sensor 62 that detects the amount of air taken in by the engine 100, and an oil temperature sensor 63 (an example of a viscosity characteristic detector). ), A cam angle sensor 64 for detecting the rotational phase of the camshafts 41 and 42 and a water temperature sensor 65 for detecting the temperature of the cooling water of the engine 100 are connected. The controller 60 obtains the engine rotation speed based on the detection signal from the crank angle sensor 61, obtains the engine load based on the detection signal from the airflow sensor 62, and determines the intake side VVT and exhaust gas based on the detection signal from the cam angle sensor 64. The operating angle of the side VVT 18 is obtained. The controller 60 is an example of a control unit.

  The controller 60 determines the operating state of the engine 100 based on various detection results, and the oil control valve 84, the first direction switching valve 96, the second direction switching valve 97, and the third direction switching valve according to the determined operating state. 98 is controlled.

  One of the controls of the controller 60 is a reduced cylinder operation. The controller 60 switches between all-cylinder operation in which combustion is performed in all cylinders and reduced-cylinder operation in which combustion in some cylinders is stopped and combustion is performed in the remaining cylinders according to the operating state of the engine 100.

  Further, the controller 60 sets the highest required hydraulic pressure among the required hydraulic pressures of the hydraulic actuator according to the operating state of the engine 100 as the target hydraulic pressure, so that the actual hydraulic pressure detected by the hydraulic sensor 50a matches the target hydraulic pressure. In addition, the oil control valve 84 is controlled to feedback control the discharge amount of the oil pump 81. As feedback control, for example, PID control can be employed. For example, the exhaust side VVT 18, the valve stop mechanisms 45d and 46d, and the oil jet 71 correspond to the hydraulic operation device.

  FIG. 6 is a graph showing a temporal transition of the duty ratio and the actual oil pressure when PID feedback control is performed during cold start. The upper diagram of FIG. 6 is a graph showing the temporal transition of the duty ratio, where the vertical axis represents the duty ratio and the horizontal axis represents the time. The lower diagram of FIG. 6 is a graph showing the temporal transition of the actual hydraulic pressure, where the vertical axis indicates the hydraulic pressure and the horizontal axis indicates the time. At time T1, engine 100 is started.

  Note that the opening degree of the oil control valve 84 increases as the duty ratio increases. In the oil pump 81, as the opening of the oil control valve 84 increases, the amount of oil flowing into the pressure chamber 81m increases and the eccentric amount of the cam ring 81d decreases. As a result, the amount of oil discharged from the oil pump 81 decreases, and the actual hydraulic pressure decreases.

  Immediately after the engine 100 is started, the oil in the oil supply passage 5 has been removed, so the oil in the pressure chamber 81m has also been removed, and the pressure chamber 81m has been decompressed. That is, at time T1, the oil control valve 84 is in a state where the oil discharge amount by the oil pump 81 cannot be controlled. Therefore, at time T1, the oil pump 81 discharges the oil at the maximum discharge amount in order to quickly spread the oil to the oil supply passage 5. As a result, the actual hydraulic pressure is rising rapidly.

  At time T2, since the actual oil pressure has greatly exceeded the target oil pressure, the controller 60 increases the duty ratio in order to reduce the actual oil pressure. Since the main gallery 50 is located upstream of the oil supply passage 5, the oil is immediately filled, and the hydraulic pressure sensor 50a detects a sudden change in hydraulic pressure. On the other hand, since the pressure chamber 81m is located downstream of the main gallery 50, even if the duty ratio is increased to increase the opening of the oil control valve 84, the oil pump 81 is not easily filled with oil. Is in a state where the oil discharge amount cannot be controlled.

  Therefore, during the period from time T2 to time T3, the actual hydraulic pressure is kept high even though the duty ratio is increased. At this time, due to the influence of the I (integral) term of PID control, the deviation between the target hydraulic pressure and the actual hydraulic pressure is integrated, and the duty ratio is larger than the target duty ratio DR aimed at. Here, the target duty ratio DR indicates a duty ratio that is assumed that the actual hydraulic pressure becomes the target hydraulic pressure.

  At time T3, the entire oil supply passage 5 is filled with oil, and finally oil is supplied to the oil pressure chamber 81m via the oil control valve 84, so that the oil pump 81 is in a state where the discharge amount can be adjusted. At this time, due to the influence of the excessive duty ratio, the actual hydraulic pressure rapidly decreases to a value that is significantly lower than the target hydraulic pressure (time T4).

  In the period from time T4 to T5, the duty ratio is gradually decreased in order to increase the actual oil pressure toward the target oil pressure. However, at the time of cold start, if the duty ratio is increased, hydraulic hunting occurs, so this inclination is made as gentle as possible. Therefore, the actual hydraulic pressure is kept low.

  In the period from time T4 to T5, the actual oil pressure is greatly reduced from the target oil pressure, and the discharge amount of the oil pump 81 becomes too small. Therefore, the oil amount in the pressure chamber 81m becomes insufficient, and the oil pump 81 is discharged. The amount cannot be adjusted. As a result, the actual hydraulic pressure is kept low.

  FIG. 7 is a graph showing the characteristics of the oil control valve 84. The vertical axis shows the flow rate of oil supplied from the oil control valve 84 to the pressure chamber 81m, and the horizontal axis shows the duty ratio.

  As shown in FIG. 7, when the duty ratio exceeds the duty ratio DA, the oil flow rate starts to increase, and thereafter, the oil flow rate increases linearly as the duty ratio increases. When the duty ratio exceeds the duty ratio DB (> DA), the oil flow rate is maintained at a constant value.

  In FIG. 7, the region on the right side of the boundary duty ratio DM indicates a dead zone where the discharge amount of the oil pump 81 does not change much even if the duty ratio is changed, and the region on the left side indicates the discharge of the oil pump 81 according to the duty ratio. It shows the sensitive band where the amount changes.

  That is, if the duty ratio exceeds the boundary duty ratio DM, the discharge amount of the oil pump 81 becomes too small, so that the oil amount in the pressure chamber 81m becomes insufficient, and the oil pump 81 cannot adjust the discharge amount. On the other hand, when the duty ratio falls below the boundary duty ratio DM, the oil discharged from the oil pump 81 supplies sufficient oil to the pressure chamber 81m, and the oil pump 81 can adjust the discharge amount. The target duty ratio DR is at a position where the duty ratio is slightly lower than the boundary duty ratio DM.

  In the period from time T4 to time T5 in FIG. 6, the oil pump 81 is controlled in the dead zone of the graph in FIG. Therefore, during this period, even if the duty ratio is lowered, the actual hydraulic pressure does not increase easily.

  Returning to FIG. At time T5, since the duty ratio has finally entered the sensitive band, the actual oil pressure has changed, but since the duty ratio is below the target duty ratio DR, the actual oil pressure is significantly larger than the target oil pressure. Has risen sharply. Therefore, at time T5, the duty ratio starts to increase in order to reduce the actual oil pressure to the target oil pressure.

  Thereafter, the actual oil pressure gradually converges to the target oil pressure while repeating hunting around the target oil pressure.

  After time T5, the actual oil pressure does not easily converge to the target oil pressure because the oil viscosity is high (hard) at the time of cold start, so the pressure loss through the oil supply passage 5 is high, and the duty ratio adjustment → oil control valve 84 This is because it takes time to repeat a series of cycles of operation → supply of oil to the pressure chamber 81m → operation of the oil pump 81 → change in the discharge amount of oil → sensing of the hydraulic sensor 50a → readjustment of the duty ratio.

  Therefore, if the oil temperature of the oil is equal to or lower than the reference oil temperature, the controller 60 sets the duty ratio (an example of the control value) of the oil control valve to a fixed duty ratio (an example of the fixed control value) for a certain period from the start of the engine 100. After executing fixed duty control (an example of fixed value control) set to (example), feedback control is executed. This eliminates the integration of the deviation due to the I term in the PID control for a certain period from the start of engine 100, thereby shortening the period from time T1 to T5 in FIG.

  FIG. 8 is a graph showing the temporal transition of the duty ratio and the actual hydraulic pressure when the fixed duty control is performed at the cold start. The upper diagram of FIG. 8 is a graph showing the temporal transition of the duty ratio, where the vertical axis represents the duty ratio and the horizontal axis represents the time. The lower diagram of FIG. 8 is a graph showing the temporal transition of the actual hydraulic pressure, where the vertical axis indicates the hydraulic pressure and the horizontal axis indicates the time.

  At time T1, engine 100 is started. At this time, the duty ratio is set to the fixed duty ratio DF, and the fixed duty control is started. Immediately after the engine 100 is started, the oil in the oil supply passage 5 has been drained, and the oil is not sufficiently supplied to the pressure chamber 81m. Therefore, as in FIG. 6, the oil pump 81 discharges oil at the maximum discharge amount. To do. As a result, the actual hydraulic pressure is rising rapidly. Thereafter, during the period up to time T2 when the fixed duty control is terminated, the actual hydraulic pressure changes in the same manner as the period of time T1 to T2 in FIG.

  At time T2 when the actual hydraulic pressure decreases to the target hydraulic pressure, the controller 60 ends the fixed duty control and starts the feedback control.

  During the period from time T2 to time T3, the actual hydraulic pressure is lower than the target hydraulic pressure, so that the duty ratio is gradually decreased to increase the actual hydraulic pressure, but the discharge amount from the oil pump 81 is also decreased due to the decrease in the actual hydraulic pressure. The amount of oil in the pressure chamber 81m becomes insufficient, and the actual hydraulic pressure remains low for a while.

  However, in FIG. 8, the controller 60 sets the duty ratio of the oil control valve 84 to the fixed duty ratio DF during the period of time T1 to T2. Therefore, during this period, the deviation between the target hydraulic pressure and the actual hydraulic pressure is not integrated, and at time T2 when the feedback control is started, the duty ratio is compared with the target duty ratio DR corresponding to the target hydraulic pressure at the start of the feedback control. Don't be oversized. Therefore, as in the case of FIG. 6, the actual oil pressure is not significantly lower than the target oil pressure, and the oil pump 81 is immediately ready to adjust the oil discharge amount (time T3). Thereafter, the actual hydraulic pressure converges to the target hydraulic pressure by feedback control, showing the same behavior after time T5 in FIG.

  As described above, in the present embodiment, since the fixed duty control is performed without performing the feedback control at the cold start, the period from the time T1 to T5 shown in FIG. 6 is the period from the time T1 to T3 shown in FIG. The actual hydraulic pressure can be quickly converged to the target hydraulic pressure.

  FIG. 9 is a flowchart showing an example of the operation of the oil supply apparatus 200 in the embodiment of the present invention. First, the controller 60 determines whether or not the oil temperature detected by the oil temperature sensor 63 is equal to or lower than a reference oil temperature (S701). As the reference oil temperature, a temperature at which the time until the actual oil pressure converges to the target oil pressure can be adopted when the fixed duty control is performed, compared to when the feedback control is performed. Specifically, 0 degrees can be employed as the reference oil temperature, but this is an example, and different values may be employed depending on the type of oil.

  Since the oil has a characteristic that the viscosity increases as the oil temperature decreases, the oil viscosity can be specified if the oil temperature is known. Therefore, this embodiment employs oil temperature as the viscosity characteristic of oil.

  Here, although the oil temperature was measured using the oil temperature sensor 63, this is an example. For example, the controller 60 may estimate the oil temperature from the cooling water temperature or the ambient temperature of the engine 100, and compare the estimated oil temperature with the reference oil temperature.

  Next, the controller 60 refers to the fixed duty ratio table and determines a fixed duty ratio corresponding to the oil temperature.

  The upper diagram of FIG. 10 is a diagram illustrating an example of the fixed duty ratio table T101. The fixed duty ratio table T101 stores the oil temperature and the fixed duty ratio in association with each other. In the example of the upper diagram of FIG. 10, five fixed duty ratios “5 degrees” corresponding to five oil temperatures of “−40 degrees”, “−30 degrees”, “−20 degrees”, “−10 degrees”, and “0 degrees”. “D0”, “D1”, “D2”, “D3”, “D4” are stored.

  Here, the fixed duty ratio indicates that the actual hydraulic pressure exceeds a predetermined upper limit hydraulic pressure when the oil control valve 84 is configured with an element having the largest variation among the variations in characteristics of the elements configuring the oil control valve 84. A value that satisfies the condition is not set. Here, since the oil control valve 84 is composed of a linear solenoid, the elements constituting the oil control valve 84 are, for example, electric circuit elements such as resistors and coils.

  As the predetermined upper limit oil pressure, when the oil pressure is further increased, it is assumed that the oil supply device 200 is damaged, oil leaks, noise noise is generated from the oil pump 81, and oil is injected from the oil jet 71. Further, it is possible to employ a value assumed that the engine 100 is cooled by the oil and the combustion stability of the engine 100 is lowered.

  Further, when the fixed duty ratio is closer to the target duty ratio corresponding to the target hydraulic pressure at the start of the feedback control, the transition from the fixed duty control to the feedback control can be performed quickly.

  Therefore, in the upper diagram of FIG. 10, the fixed duty ratio is set to a value that matches or is close to the target duty ratio corresponding to the required hydraulic pressure of the hydraulic actuator at the start of the feedback control under the above conditions. Specifically, when the target duty ratio is input to the oil control valve 84 configured with elements having the largest variation, if the actual hydraulic pressure exceeds the upper limit hydraulic pressure, the fixed duty ratio is set to the upper limit hydraulic pressure. An increased value is set to reduce to. On the other hand, when the target duty ratio is input to the oil control valve 84 configured with the element having the largest variation, if the actual hydraulic pressure does not exceed the upper limit hydraulic pressure, the fixed duty ratio is set to the target duty ratio.

  At the time of start-up, it is considered that the exhaust side VVT 18 of the hydraulic operating device operates early and the required hydraulic pressure of the exhaust side VVT 18 is the highest, so the required hydraulic pressure of the exhaust side VVT 18 is adopted as the required hydraulic pressure. The

  In the upper diagram of FIG. 10, as the oil temperature increases, the fixed duty ratio is increased, and the amount of oil discharged from the oil pump 81 is decreased. This is because the higher the oil temperature, the lower the viscosity of the oil and the better the oil response. That is, when the oil temperature is high, the oil pressure can be set to the actual oil pressure even if the amount of oil discharged from the oil pump 81 is reduced as compared with the case where the oil temperature is low.

  In addition, about the oil temperature which is not described directly in fixed duty ratio table T101, the controller 60 corresponds by linearly interpolating the fixed duty ratio of the oil temperature before and behind described in fixed duty ratio table T101. What is necessary is just to calculate the fixed duty ratio of oil temperature.

  Returning to FIG. In S703, the controller 60 refers to the duration table and determines the duration corresponding to the oil temperature.

  The lower diagram of FIG. 10 is a diagram illustrating an example of the duration table T102. The duration table T102 stores the oil temperature and the duration in association with each other. The duration is the time for which the fixed duty control is executed.

  In the example in the lower diagram of FIG. 10, five durations “T0” corresponding to five oil temperatures of “−40 degrees”, “−30 degrees”, “−20 degrees”, “−10 degrees”, and “0 degrees”. ”,“ T1 ”,“ T2 ”,“ T3 ”, and“ T4 ”are stored. The durations “T0” to “T4” are set to smaller values as the oil temperature increases. This is because the oil response is improved as the oil temperature is higher, and the actual oil pressure can be converged to the target oil pressure in a shorter time.

  Returning to FIG. In S704, the controller 60 determines whether or not the duration determined in S703 has elapsed. If the duration time has not elapsed (NO in S704), the process waits. On the other hand, if the duration has elapsed (YES in S704), the process proceeds to S705. Thereby, the fixed duty control is executed until the duration time elapses.

  In S705, the controller 60 sets a target hydraulic pressure. Here, the controller 60 sets the target oil pressure to the required oil pressure of the hydraulic actuator having the maximum required oil pressure among the hydraulic actuators. At the time of cold start, the target oil pressure of the exhaust side VVT 18 is considered to be the maximum, so the required oil pressure of the exhaust side VVT 18 is set as the target oil pressure.

  FIG. 11 is a diagram illustrating an example of a required hydraulic pressure table T601 indicating the required hydraulic pressure of the exhaust side VVT 18. As illustrated in FIG. The required hydraulic pressure table T601 stores the required hydraulic pressure of the exhaust side VVT 18 according to the oil temperature and the rotational speed of the engine 100. In FIG. 11, “Tc1”, “Tc2”, and “Tc3” indicate the oil temperature, and “500” to “6000” indicate the rotational speed of the engine 100. That is, the controller 60 reads the required oil pressure corresponding to the current oil temperature and the rotational speed of the engine 100 from the required oil pressure table T601, and sets the target oil pressure.

  Here, the controller 60 also stores a required hydraulic pressure table similar to that in FIG. 11 for other hydraulic actuators (valve stop mechanisms 45d and 46d, oil jet 71) other than the exhaust side VVT 18, and from each hydraulic required table. Then, the required oil pressure corresponding to the current operating state of the engine 100 is read. Then, the controller 60 sets the maximum required hydraulic pressure as the target hydraulic pressure among the required hydraulic pressures read from the required hydraulic pressure tables corresponding to all hydraulic actuators.

  Returning to FIG. In S706, the controller 60 performs feedback control control so that the target oil pressure set in S705 matches the actual oil pressure detected by the oil pressure sensor 50a. Thereafter, the controller 60 periodically detects the state of the engine 100 until the engine is stopped. In the detected state, the maximum required oil pressure requested by a plurality of hydraulic actuators is set as the target oil pressure. Set and perform feedback control.

  As described above, according to the oil supply apparatus 200 of the present embodiment, the fixed duty control for setting the duty ratio of the oil control valve 84 to the fixed duty ratio is performed for a certain period from the cold start. For this reason, when the PID feedback control is performed, the duty ratio is not integrated by the I term, and the actual hydraulic pressure can be quickly converged to the target hydraulic pressure even during cold start when the oil viscosity is high. As a result, early stable control of the hydraulic actuator can be achieved.

(Modification)
(1) In the flowchart of FIG. 9, the fixed duty control is executed when the oil temperature is equal to or lower than the reference oil temperature, but the present invention is not limited to this. For example, when it is obvious that the vehicle to which the oil supply apparatus is applied is used in a low temperature region, the fixed duty control may be executed unconditionally when the engine 100 is started.

  (2) In the present embodiment, the oil temperature sensor 63 is used to detect the viscosity characteristic of the oil, but the present invention is not limited to this. For example, the viscosity characteristic of the oil may be detected using a viscosity sensor that can directly detect the viscosity of the oil. In this case, in S701 of FIG. 9, if the viscosity of the oil detected by the viscosity sensor is higher (harder), the controller 60 may execute fixed duty control.

T101 Fixed duty ratio table T102 Duration table 5 Oil supply path 18 Exhaust side VVT (hydraulic actuator)
45d, 46d Valve stop mechanism (hydraulic actuator)
71 oil jet (hydraulic actuator)
50a Hydraulic pressure sensor (hydraulic pressure detector)
60 controller (control unit)
63 Oil temperature sensor (viscosity characteristic detector)
81m Pressure chamber 81 Oil pump 84 Oil control valve 100 Engine 200 Oil supply device

Claims (4)

A variable displacement oil pump,
A hydraulic actuator connected to the oil pump via an oil supply path;
A hydraulic pressure detection unit for detecting the hydraulic pressure of the oil supply path;
An oil control valve that changes a discharge amount of oil discharged from the oil pump by controlling a flow rate of oil supplied to the pressure chamber of the oil pump;
A target hydraulic pressure is set from the required hydraulic pressure of the hydraulic actuator according to the operating state of the engine, and the oil control valve is controlled so that the actual hydraulic pressure detected by the hydraulic pressure detection unit matches the target hydraulic pressure, A control unit that feedback controls the discharge amount of the oil pump,
The control unit performs the feedback control after executing a fixed value control for setting the control value of the oil control valve to a fixed control value for a certain period from the start of the engine ,
The fixed control value indicates that the actual hydraulic pressure does not exceed a predetermined upper limit hydraulic pressure when the oil control valve is configured with an element having the largest variation among the variations in characteristics of the elements constituting the oil control valve. An engine oil supply device that is set to a value that matches or is close to a control value corresponding to a required oil pressure of the hydraulic actuator at the start of the feedback control under conditions . Further comprising a viscosity characteristic detection unit for detecting the viscosity characteristic of the oil,
2. The engine oil supply device according to claim 1, wherein the control unit sets the fixed control value to be larger as the viscosity indicated by the viscosity characteristic detected by the viscosity characteristic detection unit becomes smaller. Further comprising a viscosity characteristic detection unit for detecting the viscosity characteristic of the oil,
The engine oil supply device according to claim 1 or 2, wherein the control unit sets the predetermined period to be shorter as the viscosity indicated by the viscosity characteristic detected by the viscosity characteristic detection unit decreases. Further comprising a viscosity characteristic detection unit for detecting the viscosity characteristic of the oil,
The engine oil supply device according to any one of claims 1 to 3 , wherein the control unit executes the fixed value control when the viscosity detected by the viscosity characteristic detection unit is higher than a predetermined viscosity.

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